Wilfredo Morales and Robert Fusaro

Glenn Research Center, Cleveland, Ohio

Albert Kascak

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Permanent Magnetic Bearing for 

Spacecraft Applications

NASA/TM—2003-211996/REV1

January 2008
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Wilfredo Morales and Robert Fusaro

Glenn Research Center, Cleveland, Ohio

Albert Kascak

U.S. Army Research Laboratory, Glenn Research Center, Cleveland, Ohio

Permanent Magnetic Bearing for 

Spacecraft Applications

NASA/TM—2003-211996/REV1

January 2008

National Aeronautics and

Space Administration

Glenn Research Center

Cleveland, Ohio 44135

Prepared for the

2003 Annual Meeting and Exhibition

sponsored by the Society of Tribologists and Lubrication Engineers

New York City, New York, April 28–May 1, 2003
Acknowledgments

The authors wish to acknowledge the indispensable contributions of Ben Ebihara, Ralph Jansen, and Mark Siebert, Ohio 

Aerospace Institute, Brook Park, Ohio, and Andrew Kenny, Texas A&M University, College Station, Texas.

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Level of Review: This material has been technically reviewed by technical management. 

Document Change History

This printing, numbered as NASA/TM─2003-211996/REV1, January 2008, replaces the previous version, NASA/TM─2003-211996, March 

2003, in its entirety. It contains the following changes:

Page ii, Acknowledgments text has been replaced with

The authors wish to acknowledge the indispensable contributions of Ben Ebihara, Ralph Jansen, and Mark Siebert, Ohio Aerospace Institute, 

Brook Park, Ohio, and Andrew Kenny, Texas A&M University, College Station, Texas.

Page 1, Introduction, paragraph 2: Siebert, et al. (2002) reference has been inserted so the paragraph reads

Magnetic bearings can be grouped as either active or passive. Actively controlled magnetic bearings operate in the attractive mode. They have the 

ability to support large loads, they have adjustable bearing characteristics such as stiffness and damping, and dynamic control. However, they are 

expensive and require complicated control hardware, such as digital signal processors, amplifi ers, digital-to-analog converters, analog-to-digital 

converters, and software. Passive magnetic bearings on the other hand do not require this hardware (Siebert, et al. (2002)). They can be made 

smaller, more effi cient and reliable. The disadvantages of passive magnetic bearing are that they typically have lower stiffness and lower damping than similar size active magnetic bearings, and vibration control cannot be used. 

Page 5, Present and Future Work text has been replaced with

This work was led by Robert Fusaro who was able to obtain funding from the Glenn Research Center Director’s Discretionary Fund. This work 

will serve as the basis to hopefully convince the Discretionary Fund Committee to continue further funding in the area of magnetic bearings. 

If funded, the magnetic bearing rig will be tested at higher rotational speeds, up to 50,000 rpm, to assess vibration problems which are of great 

concern if magnetic bearings are to be used in spacecraft. Another area of concern is the successful use of mechanical back-up bearings that are 

simple in design, small in size, and capable of absorbing the high rotational energy of the magnetic bearing without serious damage.

Page 6, the following reference has been inserted:

Siebert, M., Ebihara, B., Jansen, R., Fusaro, R.L., Morales, W., Kasak, A., and Kenny, A., 2002, “A Passive Magnetic Bearing Flywheel,” 

NASA TM—2002-211159, 2002.
NASA/TM—2003-211996/REV1 1 

PERMANENT MAGNETIC BEARING FOR SPACECRAFT APPLICATIONS 

Wilfredo Morales and Robert Fusaro 

National Aeronautics and Space Administration 

Glenn Research Center 

Cleveland, Ohio 44135 

Albert Kascak 

U.S. Army Research Laboratory 

Glenn Research Center 

Cleveland, Ohio 44135 

ABSTRACT 

 A permanent, totally passive magnetic bearing rig was designed, constructed, and tested. 

The suspension of the rotor was provided by two sets of radial permanent magnetic bearings 

operating in the repulsive mode. The axial support was provided by jewel bearings on both 

ends of the rotor. The rig was successfully operated to speeds of 5500 rpm using an air 

impeller. Radial and axial stiffnesses of the permanent magnetic bearings were 

experimentally measured and then compared to finite element results. The natural damping 

of the rotor was measured and a damping coefficient was calculated. 

INTRODUCTION 

 Ball bearings are used on satellites for many purposes, including positioning control and 

scanning equipment operation. Although their use in the vacuum of space has a long history, 

lubrication has remained a troublesome area especially when longer satellite lifetimes are 

demanded. Both solid lubricants and liquid lubricants have their pros and cons but neither 

one appears to minimize sufficiently the friction and wear needed for extended satellite 

lifetime. Magnetic bearings offer a desirable alternative to the use of ball bearings in a space 

vacuum. They can operate with near zero friction and wear without the use of lubricant. They 

can be used over a wide temperature range and exhibit very low vibration. 

 Magnetic bearings can be grouped as either active or passive. Actively controlled 

magnetic bearings operate in the attractive mode. They have the ability to support large 

loads, they have adjustable bearing characteristics such as stiffness and damping, and 

dynamic control. However, they are expensive and require complicated control hardware, 

such as digital signal processors, amplifiers, digital-to-analog converters, analog-to-digital 

converters, and software. Passive magnetic bearings on the other hand do not require this 

hardware (Siebert, et al. (2002)). They can be made smaller, more efficient and reliable. The 

disadvantages of passive magnetic bearing are that they typically have lower stiffness and 

lower damping than similar size active magnetic bearings, and vibration control cannot be 

used. 

For successful operation, a passive magnetic bearing must have sufficient stiffness and 

damping to perform over its entire operating range. A few studies have been published on 

the stiffness and damping of passive magnetic bearing systems. Ohji et al. (1999) 
NASA/TM—2003-211996/REV1 2 

investigated the radial disturbance attenuation characteristics of horizontal and vertical shaft 

machines supported radially by permanent magnetic bearings. Fremerey (2000) used 

permanent magnetic bearings with eddy-current damping to levitate a 500 Whr energy 

storage flywheel for an uninterruptible power supply. Satoh et al. (1996) used a mechanical 

damper attached to the stator magnets in a system with four axes that are passively levitated. 

Jansen and DiRusso (1996) measured the stiffness and damping of a ferrofluid stabilizer, and 

the stiffness of permanent magnetic bearings. 

 Jewel bearings are widely used in electrical measuring instruments such as galvanometers, 

ammeters, and energy measuring meters. A jewel bearing consists of a ball that spins on an 

axis perpendicular to the contact area. Jewel bearings have been treated analytically and 

experimentally. Sankar and Tzenov modeled the motion accuracy of a jewel bearing (1994a) 

and determined its steady-state dynamics using a free ball and its motion due to friction 

(1994b). Spinning ball contact of jewel bearings was experimentally studied at the Glenn 

Research Center in the late 60’s and early 70’s. A series of papers were written on the 

spinning ball contact. Townsend, Dietrich, and Zaretsky (1969) included speed effects on ball 

spinning torque. Allen and Zaretsky (1970) proposed a microasperity model for the 

elastohydrodynamic lubrication of a spinning ball on a flat surface. Dietrich, Parker, and 

Zaretsky (1968) studied the effect of ball-race conformity on spinning friction. Townsend 

and Zaretsky (1970) studied the effects on ball spinning torque of antiwear and extremepressure additives in a synthetic paraffinic lubricant. 

The purpose of this work was to design, construct and operate a totally passive, 

permanent magnetic bearing rig to demonstrate its potential for spacecraft applications. The 

design incorporated permanent magnetic bearings, operating in the repulsive mode for radial 

bearing support, and jewel bearings for axial support. 

EXPERIMENTAL 

DESIGN AND CONSTRUCTION 

 A photo image and a schematic view of the permanent magnetic bearing rig are shown in 

Figures 1 and 2. The support base is made of a single piece of aluminum alloy. The rest of 

the rig consists of two permanent magnetic bearings, a wheel, jewel bearings on both ends of 

the rotor, and an air impeller. The two sets of permanent magnetic bearings operating in the 

repulsion mode provide levitation in the radial direction. The axial movement of the rotor is 

constrained by two jewel bearings that were placed on both ends of the rotor to allow for 

force reversal. 

 The rotor shaft was made of nonmagnetic stainless steel. The wheel, permanent magnetic 

bearings and an air impeller were slip fit over the rotor shaft and restrained axially by 

retaining rings. Each end of the rotor has a ball retainer to hold the 3.175 mm (0.125 inch) 

diameter ball used in the jewel bearing. The ball retainer is held to the rotor by threads on the 

rotor bore. The arrangement of components on the rotor is shown in Figure 3. 

Figure 4 depicts the permanent magnet bearing assembly. The magnets were axially 

magnetized, meaning the poles are on the flat sides the magnets. A separate magnet stack 

was attached to the rotor and to the stator. Four magnet rings comprise each magnet stack. 

The poles of the magnets were oriented so that the stationary magnets and the rotating 

magnets repel each other when the non-rotating and the rotating sleeves are axially aligned. 
NASA/TM—2003-211996/REV1 3 

 Neodymium-iron-boron (NdFeB) magnets were used, having a nominal BH product of 35 

MGOe (Gauss-Oersted ×106

). The dimensions of the rotating and stationary magnets are as 

follows. The OD (outside diameter) of the rotating magnet is 45.7 mm (1.800 in) and the ID 

(inside diameter) is 31.8 mm (1.250 in). The stationary magnet OD is 61.0 mm (2.400 in) and 

the ID is 48.3 mm (1.900 in). The thickness of magnet rings used in both stacks is 5.08 mm 

(0.200 inch). Therefore, the total axial thickness of the magnets in each stack is nominally 

20.3 mm (0.800 inch). 

 Both the non-rotating and the rotating sleeves were made of nonmagnetic stainless steel. 

The stacks of magnets and the nonconductive spacers were connected together with epoxy in 

their mounting sleeve. The magnets were stacked with like poles adjacent to each other (NSSN-NS-SN) and permanently locked in place with epoxy. During the assembly process each 

magnet was coated with epoxy and put in the sleeve along with the nonconductive spacers 

and the sleeve end ring. An arbor press was used to provide the force to hold the magnets in 

the sleeve while press pins could be installed in the sleeve end ring. Once the sleeves were 

assembled, a thin coat of epoxy was also applied to the free surface of the magnets to prevent 

the magnets from oxidizing. 

 The OD of the nonrotating magnet sleeve was threaded. These threads fit the threading 

along the rotor axis of the support base, allowing the stator magnets to be moved axially. The 

stator magnet sleeve was rotated by inserting a Ø4.75mm rod into a hole on the end of this 

sleeve. The rod was turned until the stationary sleeve was positioned over the rotating sleeve. 

The stator magnets were adjusted to minimize the axial force before operation. 

 The rotating and stationary magnets had a nominal 1.27 mm (0.05 inch) radial gap. The 

nonconductive spacers on the rotor and the stator were axially aligned and have a 0.76 mm 

(0.03 inch) nominal gap. This smaller clearance was to prevent the rotor and stator magnets 

from contacting during installation and operation. 

 A schematic of the jewel bearing assembly is shown in Figure 5. The jewel bearing on 

each end of the rotor consisted of a removable 3.175 mm (0.125 inch) diameter Si3N4 ball 

held in the ball retainer in the center of the rotor shaft. About 0.50 mm (0.02 inch) of the ball 

extends beyond the surface of the ball retainer. The ball spins against the stationary 440C 

stainless steel disk of the jewel bearing. The surface of the disk was perpendicular to the axis 

of the rotor. A piezoelectric load cell measured the load of the jewel bearing contact. 

An air impeller drives the shaft. The air impeller is a disk with blind holes drilled on both 

sides. One air stream is directed to either side of the disk to minimize the radial force on the 

rotor. The blind holes catch the air, and rotate the rotor. 

SPIN TESTING 

 The rotor was spun using an air impeller. A 120-psig-air line was connected to a pressure 

regulator leading to the air impeller. Adjusting the flow rate of the compressed air against the 

impeller could control the rotational speed of the bearing. A stroboscope was used to monitor 

the rotational speed. 

 Testing was conducted by slowly increasing the speed. The speed was gradually increased 

to 5500 rpm and the bearing ran very smoothly over the entire speed range. It was stopped at 

this point only because of the loud noise generated by the compressed air. 
NASA/TM—2003-211996/REV1 4 

The rotor passed through a critical speed at 3336 rpm, but no resonant vibrations were 

observed. This critical speed was found by impact testing of the rotor and is discussed later in 

the paper. The tests were repeated a second time and again no vibrations were observed. 

RESULTS AND DISCUSSION 

RADIAL AND AXIAL STIFFNESSES OF BEARINGS 

 The procedure for experimentally measuring the radial magnetic bearing stiffness was to 

direct the axial force to one jewel bearing by adjusting the magnetic bearing stators. The 

disk part of the jewel bearing at the other end of the rotor was removed so that the rotor was 

free to rotate about the contacting jewel bearing. Two dial indicators were contacted against 

the shaft to measure the radial deflection. One dial indicator was positioned 40 mm from the 

free end of the rotor. The other dial indicator was positioned 25mm from the contacting jewel 

bearing. The rotor was deflected by adding weights to the shaft near the free end. The 

deflections of both dial indicators were recorded as a function of applied load. The moments 

about the existing jewel bearing were computed for each load applied and these were used to 

calculate the stiffness of the passive magnetic bearings. Analysis of the two dial indicator 

readings showed that the ball of the jewel bearing was not laterally displaced when the 

weights were loaded on the rotor. The force versus deflection characteristics of one passive 

magnetic bearing was calculated from the load versus deflection curve. 

To predict the stiffness of the permanent magnetic bearing by the finite element method, 

a finite element model was made for the permanent magnet rings. The model had 77760 

linear brick elements as shown in Figure 6. A magnetostatic solver was used to compute the 

distribution of the magnetic field. The Maxwell stress tensor was integrated over the surface 

of the four inner ring magnets to calculate the force on them. 

Figure 7 shows the displacement as a function of radial load for the finite element and 

experimental data. In order to compare the experimental and the finite element results, linear 

regression curve fits were performed. The slope of the experimental curve is 1.78×105

 N/m 

(1.02×103

 lbf/in). The slope of the finite element curve is 1.87×105 N/m (1.07×103 lbf/in). 

The radial stiffness found by the experimental method is about 5.2 percent less than the finite 

element prediction which was considered very good agreement. 

The radial stiffness per unit length of the magnet stack of each bearing was calculated by 

dividing the experimental radial stiffness, 1.78×105 N/m (1.02×103 lbf/in) by the length of 

each magnet stack, 20.3mm (0.800 in). This value is 8.76×106

 (N/m)/m (1.27×103 (lbf/in)/in). 

To measure the axial stiffness, one non-rotating magnetic bearing sleeve was rotated by 

means of the threading in the support base. This caused the magnet stacks to become axially 

displaced relative to each other. This axial displacement was measured with a dial indicator. 

The axial load due to this displacement was measured with the load cell positioned behind 

the disk of the jewel bearing. This procedure was repeated for a number of different 

displacements and the data plotted in Figure 8. The figure shows that the axial load is a linear 

function of displacement for this permanent magnetic bearing. A linear regression fit of the 

curve gave a value of 3.52×105

 N/m (2.01×103 lbf/in) for the axial stiffness of one bearing. 

Application of Earnshaw’s theorem implies that the axial stiffness should be exactly 

twice the radial stiffness. Comparing the experimental axial stiffness of Figure 8 [3.52×105

N/m (2.01×103

 lbf/in)] to twice the experimental radial stiffness of Figure 7 [1.78×105 N/m 
NASA/TM—2003-211996/REV1 5 

(1.02×103 lbf/in)] showed that the 2 to 1 ratio predicted by Earnshaw's theorem is within 1.1 

percent which is considered excellent agreement. 

DAMPING 

 The procedure for measuring the free vibration history of the rotor was as follows. An 

axial force on one of the jewel bearings was produced by misalignment of the stationary 

magnets and then rotating magnets of the magnetic bearing. The disk of the jewel bearing at 

the unloaded end of the rotor was removed so that the rotor was free to move about this end.. 

An accelerometer was attached to the rotor 25 mm from the free end. To excite the rotor, the 

wheel was struck radially with an impact hammer. Figure 9 gives the radial free vibration 

history of the rotor as a function of time. The rotor was “pinned” against only one jewel 

bearing since during operation only one jewel bearing is in contact. These conditions resulted 

in only one mode, not a bounce and a tilt mode as would occur with “free-free” boundary 

conditions. 

 A radial damping coefficient was calculated from the logarithmic decrement and from the 

half-power bandwidth methods. The natural logarithm of consecutive amplitude ratios is 

called the logarithmic decrement, δ, equaling ln(x1/x2), where x1 and x2 are consecutive peak 

amplitudes. The log decrement is a measure of the damping factor ξ and it gives a convenient 

method to measure the damping in a system. The damping coefficient is related to the log 

decrement, by δ=2πξ/(1-ξ2 )

0.5. Taking the log decrement of the first five peaks in the time 

domain gives ξ=0.0649, or about 6.5 percent damping. The radial frequency response of the 

rotor is shown in Figure 10. The first fundamental occurred at 55.6 Hz (3336 rpm). From the 

frequency response data, the damping coefficient of the first fundamental was calculated by 

the half-power bandwidth method. The damping coefficient of the first fundamental was 

found to be 0.069, close to the damping coefficient as found by the logarithmic decrement 

method. Since no vibrations were observed on running the bearing through the critical speed, 

6.5 percent damping is adequate to control the rotor. 

PRESENT AND FUTURE WORK 

 This work was led by Robert Fusaro who was able to obtain funding from the Glenn 

Research Center Director's Discretionary Fund. This work will serve as the basis to hopefully 

convince the Discretionary Fund Committee to continue further funding in the area of 

magnetic bearings. If funded, the magnetic bearing rig will be tested at higher rotational 

speeds, up to 50,000 rpm, to assess vibration problems which are of great concern if 

magnetic bearings are to be used in spacecraft. Another area of concern is the successful use 

of mechanical back-up bearings that are simple in design, small in size, and capable of 

absorbing the high rotational energy of the magnetic bearing without serious damage. 

SUMMARY OF RESULTS 

A totally passive, permanent magnetic bearing rig was designed, constructed, and tested. 

Support of the rotor was accomplished in the radial direction by two sets of permanent 

magnetic bearings and in the axial direction by two sets of jewel bearings. The following 

results were obtained. 
NASA/TM—2003-211996/REV1 6 

1. The rig was successfully operated to a speed of 5500 rpm, which was 65 percent 

above the first critical speed. 

2. The radial stiffness of one magnetic bearing was measured to be 1.78×105 N/m 

(1.02×103 lbf/in) which was within 5 percent of a finite element prediction. The 

measured stiffness was adequate for the currently designed rotor. 

3. The fact that experimental and finite element methods predictions agree demonstrate 

that finite element methods can be used to accurately design the size and number of 

the magnets for the radial support of the bearings. The radial damping for the first 

fundamental was measured to be 6.5 percent of critical damping. This was adequate 

damping to pass through the first critical speed. 

4. The axial stiffness was measured and was shown to be within 1 percent of twice the 

measured radial stiffness. This is consistent with Earnshaw’s theorem. 

 REFERENCES 

Allen, C.W., Zaretsky, E.V., 1970, “Microasperity Model for Elastohydrodynamic 

Lubrication of a Spinning Ball on a Flat Surface,” NASA TN D–6009. 

 Dietrich, M.W., Parker, R.J., Zaretsky, E.V., 1968, “Effect of Ball-Race Conformity on 

Spinning Friction,” NASA TN D–4669. 

Fremerey, J.K., 2000, “A 500-Wh Power Flywheel on Permanent Magnet Bearings,” 

Fifth International Symposium on Magnetic Suspension Technology, pp. 287–295. 

 Jansen, R., DiRusso, E., 1996, “Passive Magnetic Bearing with Ferrofluid Stabilization,” 

NASA TM–107154. 

Ohji, T., Mukhopadhyay, S.C., Iwahara, and Yamada, S., 1999, “Permanent Magnet 

Bearings for Horizontal- and Vertical-Shaft Machines: A Comparative Study,” Journal of 

Applied Physics, Vol. 31, 8, pp. 4648–4650. 

 Sankar, T.S., and Tzenov, P.I., 1994a, “Friction and Motion Accuracy in Jewel Bearing,” 

Tribology Transactions, Vol. 37, 2, pp. 269–276. 

Sankar, T.S., and Tzenov, P.I., 1994b, “The Steady-State Dynamics of Jewel Bearing 

with a Free Ball,” Tribology Transactions, Vol. 37, 2, pp. 403–409. 

Satoh, I., Shirao, Y., and Kanemitsu, Y., 1996, “Dynamics and Vibration of a Single Axis 

Active Magnetic Bearing System for Small-Sized Rotating Machinery,” Fifth International 

Symposium on Magnetic Bearings, pp. 497–502. 

Siebert, M., Ebihara, B., Jansen, R., Fusaro, R.L., Morales, W., Kasak, A., and Kenny, 

A., 2002, “A Passive Magnetic Bearing Flywheel,” NASA TM⎯2002-211159, 2002. 

Townsend, D.T., and Zaretsky, E.V., 1970, “Effects of Antiwear and Extreme-Pressure 

Additives in a Synthetic Paraffinic Lubricant on Ball Spinning Torque,” NASA TN D–5820. 

Townsend, D.P., Dietrich, M.W., Zaretsky, E.V., 1969, “Speed Effects on Ball Spinning 

Torque,” NASA TN D–5527. 

 
NASA/TM—2003-211996/REV1 7 

Figure 1.—Photographic image of the magnetic bearing rig.

Figure 2.—Schematic view of magnetic bearing rig. The permanent

 magnetic bearings and jewel bearing assembly at the near end 

 were sectioned for clarity.

Permanent magnetic bearing

Air impeller

Permanent Wheel

magnetic bearing

Jewel

bearing

Support

base
NASA/TM—2003-211996/REV1 8 

Figure 3.—Schematic of rotor that shows the location of 

 components. The total mass of the rotor is 2.26 kg. The

 diameter of the wheel is 102 mm.

0 58 mm 110 mm 192 mm 263 mm 320 mm

Mag 

bearing

rotor

Mag bearing

rotor

Air

impeller

Rotor shaft

Wheel

Ball of jewel bearing

Figure 4.—Sectioned view of permanent magnetic bearing assembly.

 There are two of these assemblies attached to the rotor.

Nonrotating 

sleeve

Rotating 

sleeve

Nonconductive 

spacer

Nonconductive 

spacer

Rotating

magnets

Stationary

magnets

Sleeve 

end ring

Sleeve 

end ring
NASA/TM—2003-211996/REV1 9 

Figure 5.—Jewel bearing assembly applied to each end of the 

 rotor. The ball of the jewel bearing rotates against the stationary

 disk of the jewel bearing during operation.

Load

cell

Ball of jewel

Disk of jewel bearing

bearing

Ball 

retainer

Conical backup

surface

Shaft of

rotor

Threaded rod

Figure 6.—Finite element model of ring magnets

 in magnetic bearing.
NASA/TM—2003-211996/REV1 10 

Figure 7.—Radial displacement as a function of radial load for one permanent magnetic

 bearing as determined by experimental testing and finite element analysis.

Experimental

Finite element

120

100

80

60

40

20

0

0.0 0.1 0.2 0.3

Radial displacement, mm

0.4 0.5 0.6 0.7

Radial load, N

Figure 8.—Axial displacement as a function of axial load for one magnetic-bearing stator

 of the permanent magnetic bearing.

600

500

400

300

200

100

0

0.0 0.2 0.4 0.6

Axial displacement of one bearing stator, mm

0.8 1.0 1.4 1.8 1.2 1.6

Axial load, N
NASA/TM—2003-211996/REV1 11 

Figure 9.—Radial free vibration history of the rotor from an impact test.The vertical axis

 is the output of the accelerometer attached to the free end of the rotor.

1.00

0.75

0.90

0.60

0.30

0.45

0.15

0.00

–1.00

–0.30

–0.15

–0.60

–0.45

–0.90

–0.75

0.00

Time, sec

0.05 0.10 0.15 0.20 0.25 0.30 0.35 0.40 0.45 0.50

Accelerometer output

Figure 10.—Frequency response of the rotor that occurred from impact test.

6.4

7.2

5.6

4.0

4.8

3.2

2.4

0.8

0.0

1.6

0

Frequency, Hz

10 20 30 40 50 60 70 80

Amplitude
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Permanent Magnetic Bearing for Spacecraft Applications 

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Morales, Wilfredo; Fusaro, Robert; Kascak, Albert 

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13. SUPPLEMENTARY NOTES

14. ABSTRACT

A permanent, totally passive magnetic bearing rig was designed, constructed, and tested. The suspension of the rotor was provided by two 

sets of radial permanent magnetic bearings operating in the repulsive mode. The axial support was provided by jewel bearings on both ends 

of the rotor. The rig was successfully operated to speeds of 5500 rpm using an air impeller. Radial and axial stiffnesses of the permanent 

magnetic bearings were experimentally measured and then compared to finite element results. The natural damping of the rotor was 

measured and a damping coefficient was calculated. 

15. SUBJECT TERMS

Magnetic bearings; Jewel bearings; Space vehicles; Aerospace; Spacecraft 

16. SECURITY CLASSIFICATION OF: 17. LIMITATION OF 

 ABSTRACT

UU 

18. NUMBER 

 OF 

 PAGES

17 

19a. NAME OF RESPONSIBLE PERSON

STI Help Desk (email:help@sti.nasa.gov) 

a. REPORT

U 

b. ABSTRACT

U 

c. THIS 

PAGE

U 

19b. TELEPHONE NUMBER (include area code)

301-621-0390 

Standard Form 298 (Rev. 8-98)

Prescribed by ANSI Std. Z39-18